This invention relates generally to positive displacement pumps and more particularly to sealless screw pump/motor packages especially for pumping multi-phase fluids in subsea applications.
As remote subsea wells deplete, boosting is not cost effective if the pump requires mostly liquid in order to function; because such wells produce a large volume fraction of dirty water and gas along with a small amount of oil. The small amounts of oil involved, 1000 barrels per day (bpd) or less, cannot be economically recovered unless a multiphase pump is located in the vicinity of the well. To improve economics, multiple wells can be manifolded together to feed a single pump, the piping arrangement providing for a flow check of each pump individually. This concept is illustrated in FIG. 1. Several of such multiphase pumps delivering product to centrally located separation equipment on a surface platform or onshore appears to be a practical way to extend the life of wells that would otherwise have to be abandoned. These wells normally produce mixtures of gas, oil and water in varying proportions that can vary considerably at the pump inlet over time. Gas void fractions (GVF) of 0.95 (i.e., 95% gas by volume)xe2x80x94and higherxe2x80x94are fairly typical. GVF is related to the more frequently quoted gas-oil ratio (GOR) or the mass of gas in standard cubic feet per barrel of oil (scf/bbl) as follows:
GVF=GLR/(1+GLR)xe2x80x83xe2x80x83(1)
where GLR is the volume flowrate ratio of gas QG to liquid QL and is given by
GLR=(GOR)(T/Tstd)(Pstd/P)/(5.615 cu ft per bbl)xe2x80x83xe2x80x83(2)
where T is absolute temperature and p is pressure. Standard temperature and pressure are 15xc2x0 C. and 14.7 psia respectively; so that Tstd=(273.15+15)xc2x0K. This mixture must be pumped to as much as 50 bar or 700 psi.
To date, practically all multiphase pumps have been located on the surface and generally onshore, where the installation costs are smaller and the frequent maintenance needed for new concepts can be carried out with relative ease. To install and maintain a pump subsea requires a considerable infusion of deepwater technology, which is as sophisticated as the design of the pump package itself. As more success is achieved in dealing with the technical and reliability issues encountered in the multiphase pumps located on the surface, there is now more impetus to place them subsea.
For pumping multiphase fluids, two quite different types of multiphase pump are employed, namely, a) rotodynamic and b) rotary positive displacement. Type (a) creates pressure dynamically; i.e., shaft torque is converted into fluid angular momentum. The pressure rise then depends on the product of average fluid density and velocity change. The helico-axial configuration is the rotodynamic concept that is used for multiphase pumping, because it has many axial-flow stages that do not vapor-lock; i.e., they do not separate the gas and liquid phases by the centrifugingxe2x80x94as can occur, e.g., in a single-stage centrifugal pump (also a rotodynamic machine). This machine depends on speed and fluid density to develop pressure. Sudden changes in fluid density, as would occur in slugging, produce sudden changes in torque. Type (b) develops pressure hydrostatically and so does not depend on the pump speed or fluid density. The inlet of the pump is walled off from the discharge, e.g. in the case of the popular two-screw configuration, by the meshing of the screws. As with a reciprocating pump, the shaft power is simply the displacement volume rate Qd times the pressure difference xcex94p across the pump; and the shaft torque is this power divided by the angular speed xcfx89 of the drive shaft. Thus if slugging occurs and the xcex94p remains constant, this slugging has a relatively small effect on shaft torque.
In both cases, the intake volume flowrate capability increases with speed. A rotodynamic pump needs to speed up at high GVF (low average fluid density) in order to maintain xcex94p at the same level that a lower speed produces at lower GVF; while a positive displacement pump can run at constant speed; albeit with reduced liquid output.
The efficiency of multistage pumping is the ideal power Pi divided by the pump shaft power Ps. In the presence of typical amounts of liquid, the process tends to be isothermal, in which case Pi=Pisoth, where
Pisoth=mRT11n(p2/p1)+QLxcex94pxe2x80x83xe2x80x83(3)
whereas, for no liquid flow QL present, the process tends to be adiabatic, in which case Pi=Pad, where
Pad=mcpJT1{[p2/p1]exp[xcex3xe2x88x921)/xcex3]xe2x88x921}xe2x80x83xe2x80x83(4)
In these equations, m is the mass flowrate, R is the gas constant, cp is the specific heat of the gas at constant pressure, J is the mechanical equivalent of heat, xcex3 is the ratio of specific heats of the gas, and subscripts 1 and 2 denote pump inlet and discharge respectively.
Multistaging minimizes the shaft power for a given ideal power, especially for high pressure ratios p2/p1. Such multistaging is necessary for helico-axial pumps to work; however, a single stage is the normal embodiment of a screw pump. Screw pumps tend to be smaller; so that efficiency may not then be an issue. In view of this, screw pumps are preferable for subsea applications because the small sizes needed for the low flowing remote wells are relatively inexpensive. Further economies are to be had in that they can be driven subsea by correspondingly small, constant-speed, submersible electric motors; thereby eliminating the need for VFD""s or subsea deployment of hydraulic lines to run variable-speed turbines. Also, torque shock does not occur with slugging, thereby simplifying the mechanical design of the rotors.
The mechanical design of a two-screw pump is relatively simple, because a double-suction configuration is utilized. Each rotor ingests the fluid from both ends and conveys it to the center, where it is discharged, providing an axial balance that insures long bearing life. The screws do not touch each other, and clearance is provided between the screws and the surrounding bores in the body. The two rotors are kept clear of each other by a set of timing gears that are lubricated by clean oil, along with the adjacent bearings, seals being required to isolate this oil from the pumpage. The total diametral clearances and those between the meshing screw threads do not vary with axial position; so, when pumping 100% liquid, the leakage across each land and through each portion of the mesh is the same and produces a linear development of pressure vs. axial length.
Multiphase pumps depend on the liquid sealing of these clearances to produce a net positive flowrate vs. what would otherwise be a massive leakage from discharge back to inlet. In the case of 100% gas (GVF=1), this liquid sealing is maintained by recirculating liquid that was previously captured by a phase-separation plate in the discharge zone at the center of the rotors. The reservoir for this captured liquid is the special feature of a multiphase screw pump that makes possible sustained operation at GVF=1 and which can be seen in FIG. 2. In fact, the liquid sealing is so effective at high GVF that no gas leaks back to the inlet or suction cavities at the ends of the screws. This is illustrated in the laboratory test data of FIG. 3 for the total intake volume flowrate Q1 vs. the pressure difference xcex94p across the pump. Except for the very small leakage of sealing liquid, usually less than 1% of Q1, the volumetric efficiency xcex7v, where xcex7v=Q1/Qd, is therefore 100%.
The development of pressure along the screws at high GVF is not linear with axial position as it would be for pure liquid (GVF=0). This is because gas leaks (along with the sealing liquid) across the higher-pressure screw lands near the center of the pump in order to compress the gas in the neighboring. xe2x80x9clockxe2x80x9d or trapped volume between successive mesh points along the length. The pressure drop across the last one or more lands at each of the outer, low-pressure ends of the rotors, is quite smallxe2x80x94just enough to maintain a liquid seal, so that the gas doesn""t blow back to the pump inlet. So, the pressure develops slowly at the inlet ends of the screws and more rapidly closer to the center (the pump discharge). The number of locks must therefore be sufficient to prevent blow-back. Analysis of this two-phase leakage across the lands shows that the number of locks must increase with pump xcex94p, rotor diameter, internal clearance, and decreasing sealing leakage. (There is one more land than the number of locks.) The resulting axial pressure distribution produces a radial load on the screw rotors, which is a consequence of the helical screw pitch. Screw pitch is defined by the specific pumping requirements and takes into account the xcex94p, viscosity, and required flowrate. The higher the pitch for a given diameterxe2x80x94or the higher the helix angle, the greater the load. One way to reduce this load is to reverse the direction of flow in the screws so that the fluid enters at the center and flows outward to the discharge, which is now at the ends of the screws and puts discharge pressure on the adjacent seal faces. Finite-element stress analysis reveals that, unfortunately, only a small reduction in radial displacement can be realized by this reversal of flow direction. A stronger rotor is perhaps the best approach, as pressure is kept off the seals and the rotor is more robust.
Consideration of all these factors allows the development of a full range of multiphase two-screw pumps, some having quite high flowrates, as shown in Table 1.
Capacities shown are for GVF of 0.90 with liquid viscosity of 10 cp and are approximate for general sizing purposes. Specific performance data are calculated for each application for the pump size and screw pitch.
Divide by 6.3 to get m3/d
Displacement volume rate Qd is a function of the screw rotor tip diameter D, typical values of which are found from
D=K(Qd/N)⅓xe2x80x83xe2x80x83(5)
where K ranges from 4 to 7, depending on xcex94p. D is in inches, Qd in m3/day, and rotative speed N in rpm.
For subsea applications, the screw pumps described herein are configured with submersible motors for integration of pump and motor into a viable subsea package. These may be three-phase squirrel cage wet motors with power levels ranging from 1 to 5000 kW and speeds from 200 to 3500 rpmxe2x80x94at voltages up to 10,000 V. Besides standard applications, such motors have been used for special applications in offshore, cavern, and subsea environments.
There are three basic configurations of submersible motors for subsea applications; namely, a) standard, water-filled motor, b) oil-filled motor, and c) canned motor.
Water-filled motors are widely used in submersible applications. The liquid is either water or water/glycol, which both lubricates the bearings and cools the motor. Cooling is very effective, so that additional cooling devices are not needed. The winding wire used is insulated with PVC or PE, which tightens against the high pressure. These motors have high reliability and durability.
Oil-filled motors have the same high reliability as do the above water-filled motors but are somewhat larger. A special oil-protected wire is used for the windings. An oil-filled motor is preferred for the subsea multiphase applications discussed herein. It is close-coupled to the pump, so that the oil also lubricates the timing gears and inboard bearing of the pump. A pressure compensating system maintains the oil pressure at a pressure slightly greater than that of the pump suction. Therefore, the motor case must have sufficiently thick walls to withstand well shut-in pressures (up to 350 bar). Adequate cooling can be had to the surrounding seawater by the provision of fins or coils, as needed, to facilitate the needed heat transfer.
Canned motors are used where the liquid would be corrosive to the windings and/or injurious to the insulation. They have a very thin covering of sheet metal (the can) between the stator and the rotor. The stator is filled with a special resin material for insulation, and this material requires special provisions for cooling. The thin can makes these motors vulnerable to the passage of foreign particles between the rotor and stator.
These submersible motors have been used since the late sixties in dredging and offshore working vehicles. They are driving hydraulic power packs, dredge pumps, tracking wheels, elevators, cutters, etc. The subsea vehicles are controlled through an umbilical from a support vessel on the surface. Speed control is possible by varying the speed of motor-generator sets on board the support barge. Subsea application has resulted in only minor changes to the basic design of these submersible motors. A recent example is a trenching system, which includes five 220 kW submersible motors at 60 Hz and 6600 V.
A number of rotary two screw pumps have been used in applications involving multiphase products over the last 30 years. Multiphase screw pumps have been used in the chemical processing, pulp and paper and petrochemical industries. In the last decade, the multiphase pumping applications have concentrated on petroleum products, specifically oil wells. The majority of these applications are surface located and generally onshore. One such application is located in a remote area of Alberta, in western Canada. This relatively small multiphase screw pump is connected to a field of approximately 50 small oil wells. The pump was designed to operate at a GVF of 0.663. This and the other design conditions of service are given in Table 2 and provide for ingestion of 983 m3/day of gas together with a total liquid capacity of 500 m3/day or 3145 bpdxe2x80x94at a pressure rise of 300 psi (21 bar).
The pump is equipped with a special cast body with an integral liquid separating chamber. As the multiphase mixture exits the screw area it must pass though the separating chamber where the fluid velocity is reduced, thus allowing the liquid component to separate from the gas and settle into a chamber under the screw bores. The liquid is then recirculated through cyclone separators and fed back to the inlet areas by way of the mechanical seals. This provides cooling and lubricating liquid for the seals as well as sealing liquid for the pumping screws and allows the pump to operate at GVF values of 1 (i.e., 100% gas) for extended time periods, as long as there is some recirculating liquid available to provide the sealing and cooling,required.
The pump is installed downstream of the free water knockout tank and the speed is controlled to reduce the pressure in this tank from the original 200 PSIG to 40-50 PSIG. With the present wells, this is accomplished with the pump operating at only 60-90% of full speed. The actual GVF of the product varies from 0.75 to 0.90. The benefits are listed in Table 3 and include an 8% increase in oil production with no increase in power draw and a reduction in the system pressure upstream of the pump.
No increase in power required
Reduction in well head and system pressure
Reduced differential pressure on downhole progressive cavity pumps
Reduced maintenance of downhole PC pumps: estimated two times service life
Well production capacity increased by 8%
Greatly reduced system maintenance costs
Another significant benefit is a greatly increased maintenance life of the downhole progressive cavity pumps. The reduced differential pressure on these small pumps has significantly reduced the wear and they are experiencing approximately 2 times the normal life for such pumps. This has significantly reduced the maintenance costs involved with pulling the pumps from the wells when service is required.
This field experience has confirmed that the weakest area of this multiphase pump is the mechanical seals. The modified flushing system ensures there is a source of clean flush available to cool and lubricate the seals without requiring an external flush system. This makes the pump suitable for remote locations where no separate flush system is available.
The subsea version of the multiphase screw pump has a number of design changes to allow submerged operation at high ambient pressures. FIG. 6 shows an MP1-150 size screw pump connected to a submersible liquid cooled motor. As indicated in Table 4, the unit is sized to ingest 2932 m3/d (18,440 bpd) of liquid and gas and to increase the pressure by 30 bar.
At a nominal GVF of 0.95, the unit consumes 150 hp. This power level increases to 177 hp when pumping 200 cp liquid. The design incorporates a high-pressure fabricated screw pump body with a replaceable cast liner. This design provides an integral liquid separator, which separates the liquid and provides a reservoir at the lower area of the body to store the separated liquid. This separated liquid is recirculated back into the suction areas of the screws to provide the required sealing liquid at very high GVF""s.
The cast liner portion of the body contains the precision ground bores where the screws operate with controlled clearances. The drawing shows. O-ring type sealing joints between the liner and the body, which are suitable for pressures up to approximately 2000 psi. For applications above this pressure, different gasketted joint designs are possible to permit this pump to handle high differential static pressures, which could be encountered in a deep subsea application.
This design utilizes two mechanical seals at the inboard end of the pump to seal the product from the lube oil cavity. The seals operate in the lube oil, which provides lubrication and cooling. The front mounted timing gears and thrust bearings are also mounted in the same lube oil chamber which is also connected to the submersible oil filled motor. The wall sections and sealing joints are presently designed for the 2000 psi operating pressure but can be redesigned to handle higher pressures for deeper well applications.
A differential pressure compensator is connected between the pump inlet and the lube oil chamber to control the differential operating pressure on the mechanical seals. The differential pressure compensator shown, utilizes an internal piston mechanism to regulate the differential pressure across the seals to 10% of the pump suction pressure. Other types of pressure compensators can also be used to maintain a constant differential pressure across the seals. The compensator also provides a reservoir of lube oil to make up for minor seal leakage. The sizing and type of compensator are dependent on the seal design and operating conditions and would be sized to provide adequate seal life in subsea applications.
The line bearings at the outboard end of the pump are designed as product-lubricated sleeve bearings. These silicon carbide bearings are capable of supporting the shaft loads and operating in the liquid available. A separate flush porting arrangement, not shown will direct the separated liquid in the reservoir to the ends of these bearings to provide suitable lubrication.
Temperature monitors can be installed to shut down the pump in conditions of high seal or bearing temperatures. This elevated temperature condition would occur only if a slug of 100% gas, of significant duration, were pumped. At GVF=1, the recirculated liquid will eventually dissipate with the gas and insufficient cooling will be available for the bearings. Similarly, if the lube oil supply is lost, high seal temperature will provide a warning for shutdown prior to failure. The pump can be readily restarted when some cooling liquid is available from the product.
The foregoing illustrates limitations known to exist in present subsea multiphase pumps. Thus, it would clearly be advantageous to provide an alternative directed to overcoming one or more of the limitations set forth above. Accordingly, a suitable alternative is provided including features more fully disclosed hereinafter.
In one aspect of the present invention, this is accomplished by providing a pump, including a motor and a pump housing, for pumping mixed gas and liquid, said pump comprising two intermeshed screw members for providing progressive cavities for transporting mixed fluids, within a pumping cavity, from a suction passage to a discharge reservoir of the pump housing; and means for providing cooling and lubrication to the motor and to bearings and timing gears of the screw members.